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    Pump, Hydraulic Description

    All hydraulic systems require a source of hydraulic power. In most applications, the source of hydraulicpower is a variable delivery axial piston pump. Variable displacement means that pump outlet flowvaries according to system flow demands (as more sources actuator, motors, etc. are operating,the pump will increase output to maintain maximum pump outlet pressure). Other sources of hydraulic

    power are vane or gear pumps (see Motors, Hydraulic Descriptionfor description of vane or gearrotating pumps/motors), or an accumulator (see Accumulators, Hydraulic - Description). An aerospacevehicles main hydraulic pumps are usually mounted on the engine and connected to the enginerotating shaft through a gearbox. Pumps may also be driven by an electric motor, APU, ram airturbine, or second hydraulic system (using a hydraulic motor and pump combination).

    To understand how a variable delivery axial piston pump operates refer to the pump cross sectionalview shown in Figure 1. The figure shows pump components and also how pump outlet pressure iscontrolled through a compensator valve and control piston arrangement. The key element in control ofpump outlet flow is control of the swashplate angle, , which in turn controls piston relativedisplacement and hence pump flow. Ideally, the pump delivers zero flow when there are no flowdemands and the required flow when required while maintaining system design pressure at all times.Variable displacement pumps dont obtain this ideal goal, but they do come close when flow demands

    are within their design flow range.

    Prior to the introduction of variable delivery pumps 50-60 years ago, pumps were fixed delivery whichmeans they always delivered the same amount of fluid irregardless of system flow needs. Unusedfluid was then ported back to the reservoir though a pressure relief valve. Fixed delivery pumpswasted a lot of energy through heat. A fixed delivery pump has a fixed swashplate angle and thereforeno compensator valve or control piston (refer to Figure 1). Fixed delivery pumps may be found todayin very specialized applications where the fixed delivery (or flow) is tailored to the application tominimize wasted energy. The benefit in this case is a lighter, less expensive pump. An example mightbe a standalone hydraulic system to power a large cargo door where flow demands are constant whilethe door is moving and the pump would only be on during door movement. A pressure relief valveshould also be installed in this application.

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    Figure 1 Simplified Pump Schematic

    Referring again to Figure 1, pump flow rate is determined by the swashplate angle, ,which in turncontrols pump piston displacement. Swashplate angle is controlled by the compensator valve and

    control piston. The compensator valve sets the no flow outlet pressure of the pump (e.g., 3000 50psi) and meters flow to the compensator piston. The compensator is fed hydraulic pressure from thepump outlet and is positioned based on a force balance between compensator chamber pressureacting on the piston and spring force (plus, to a lesser extent, friction and flow forces). The housing isfixed and does not rotate. The housing must remain fixed so that inlet/outlet ports, compensator valve,solenoids, and other equipment can be mounted to the pump. All other parts rotate at the pumpspeed.

    At the compensator pressure setting (e.g., 3000 50 psi for a 3000 psi system), the swashplate angle,, is zero. As the pump flow increases, the pump outlet pressure decreases. As the o utlet pressuredecreases, the compensator moves towards the closed position and the pressure on the control pistonis reduced. As the control piston pressure is reduced, the control piston spring pushes the controlpiston so that the swashplate angle, , increases, resulting in greater piston stoke and increased flow


    Figure 2 shows 3-dimensional view of a piston hydraulic pump. This figure provides a better view ofhow a pump is built and how the pistons and swashplate operate. Not shown in the figure is acompensator valve and control piston, so this is more representative of a fixed delivery pump. Thepistons are attached to the swashplate via a spherical bearing arrangement. The swashplate does notrotate. As the cylinder rotates the piston is at the lower end of the cylinder during one part of arevolution (intake) and at the top end of the cylinder during the other of the rotation (high pressureoutlet side of the pump). During 1 revolution of the swashplate, each piston will pull in fluid and pushout high pressure fluid once. For a nine piston pump, this will lead to 9 pressure pulsations per 1revolution of the swashplate.

    Figure 2 Pump Cross Section

    The relationship of outlet flow to outlet pressure for a variable delivery pump is shown in Figure 3. Thisplot can be used to estimate pump flow for a given outlet pressure. The plot also shows the flow ratewhere pump outlet pressure starts to drop off dramatically. It is important to note that Figure 3represents pump characteristics for a fixed pump rotational speed (RPM) or displacement (in


    Normally these curves are provided for the rated pump speed, but in aircraft engine speeds vary -hence pump speed varies and maximum flow varies also. Therefore, the curve shown in Figure 3 will

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    shift down for lower pump speeds and shift up for higher pump speeds up to the maximum flow of thepump.

    Figure 3 Typical Flow vs Outlet Pressure Plot (for a given pump rotational speed)

    The response of the pump to a change in outlet flow is on the order of 50 milliseconds.

    Figure 4 shows the relationship between pump flow, efficiency and outlet pressure. The drop off inflow occurs due to hydraulic fluid leakage at higher pump outlet pressures (higher delta pressureacross the piston seals), which is equivalent to volumetric efficiency. As volumetric efficiency drops off,

    pump outlet flow drops by the same amount. The other curve in Figure 4 is the overall efficiency curve(overall efficiency = volumetric efficiency x mechanical efficiency). Pump horsepower increaseslinearly with pump speed.

    Figure 4 Typical Performance Curves for a Variable Delivery Piston Pump

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    Pump Design Considerations

    The most important characteristics for a hydraulic pump are listed below. These parameters assume avariable delivery constant pressure pump.

    Rated Pressure this is the nominal pressure setting of the pump and must be compatible with thedesign operating pressure for the system (e.g., 3000 50 psi for 3000 psi system or 5000 50 psi fora 5000 nominal psi system).

    Rated Speed this is the nominal speed rating of the pump. The gearbox connecting the pump to thedriver (engine, APU, etc) will need to be compatible with the drive unit speed and the rated pumpspeed. The minimum RPM of the pump may also need to be considered

    Design Displacement this is the flow per revolution of the pump (in3/rev) that the pump is capable of

    achieving without a significant reduction in outlet pressure.

    Flow vs Pressure Curve this is a plot that has flow rate on the y-axis and pump outlet pressure onthe x-axis. This plot shows the drop off in pressure at a given flow conditions and shows where theknee in the flow curve lies (see Figure 4). This graph is required for a simulation model. Nominally, thisgraph is provided at the rated speed of the pump. If available, this graph for various pump speedswould be helpful - otherwise the flow can be ratioed using the design displacement for different pumpspeeds.

    Temperature Rating the pump must be rated for the temperature extremes that it will see inoperation, such as engine nacelles. Pump seals are the most critical component when consideringtemperature.

    Case Drain Pressure this is the nominal pressure that would be in the pump case. All pumps have acase drain line to provide a flow path to the reservoir for hydraulic fluid that flows by the piston seals

    and fluid that flows through the compensator. Without a case drain the pressure would blow out acase or shaft seal. The case drain pressure needs to be greater than the reservoir pressure (and lineresistance) to assure drainage from the case to the reservoir.

    Inlet Line Size Standard pumps will have an inlet port sized by the manufacturer. The connectinginlet line/hose will need to be compatible with the port size and type.

    Outlet Line Size- Standard pumps will have an outlet port sized by the manufacturer. The connectingoutlet line/hose will need to be compatible with the port size and type.

    Case Drain Line Size- Standard pumps will have a case drain port sized by the manufacturer. Theconnecting line/hose will need to be compatible with the port size and type.

    Recommended Inlet Pressure to operate properly a pump must be supplied sufficient hydraulic fluidat a pressure level sufficient to fill the piston cylinders as the pump rotates. A pump manufacturer willprovide recommended inlet pressures and the reservoir and reservoir to pump hydraulic lines need todesigned/sized to meet this requirement. In some instances, a boost pump may be required toachieve desired inlet pressures and flows.

    Number of Pump Pistons most aircraft piston pumps have 9 pistons. An odd number of pistons havebeen shown to have smaller output pressure fluctuations than an even number of pistons andexperience has shown 9 pistons to be an optimum number for performance.

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    Power Requirementswhat horsepower is required to drive the pump at its maximum operatingcondition. Horsepower is the product of flow and delta pressure across the pump, divided by theoverall efficiency of the pump.

    Maximum Fluid Viscosity Pump manufacturers will provide a recommendation for the maximumrecommended fluid viscosity. If fluid viscosity is greater than the recommended value, then the pump

    may start to cavitate with the pump inlet at the recommended inlet pressure. Maintaining the fluidwithin the necessary viscosity range will affect the pump inlet system design.

    Seals Seals must be compatible with the specific type of hydraulic fluid used in the pump.Specifically, seal material for carbon based fluids and synthetic fluids are different.

    Filtration Requirements A pump manufacturer will provide recommended fluid cleanlinessrequirements to help ensure reliable pump operation. This filtration requirement needs to be taken intoconsideration when designing/sizing the return filters in the hydraulic system.

    Weight Is always a concern on aircraft. A oversized pump is not desirable from both a weight andcost standpoint.

    Envelope engine nacelles and APU installations usually have limited space available and the pumpinstallation must be compatible with available volumes

    Shaft Type Shafts are usually splined and the spline characteristics must be defined to ensure aproper interface to the

    Direction of Rotation Pumps can rotate either clockwise or counterclockwise and may be of aconcern in certain installations.

    Mounting The bolt flange mounting of the pump must be compatible with the attachment on thegearbox or other attaching plate.

    Relief Valve A relief valve should be installed in the outlet (high pressure) line downstream of thepump. This relief valve will provide protection against hydraulic shock loads, thermal expansion andany possible overpressure condition. The relief valve setting should be 5-10% greater than maximumpump pressure.

    Pump Installation Considerations

    Considerations for the mounting/installation of pumps include vibration, temperature, alignment ofdrive motor to pump, spline matching and torque requirements. In some applications, a means ofquick installation and removal is required. Quick installation devices must uphold rigidity of the pumpinstallation and maintain alignment of rotating shaft.

    Vibration Need to consider vibration from power source such as the engine or APU, vibrationcharacteristics from the pump, g-loading and possibly flutter. Mounting should be sufficient towithstand these loads from both a stress and fatigue standpoint. Testing to appropriate levels fromRTCA-DO160 or MIL-STD-810, Method 510 should be conducted.

    Temperature Due to the high speed and compression of fluid, pumps operate at high temperatures.Additionally, the power source for the pump, such as the engine, is at high temperature. Temperature

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    considerations should include pump seals, fluid temperature, mounting pads, connecting hoses ortubes, etc.

    Pump/Motor Alignment Alignment of pump to drive motor splines needs to held to tight tolerances.Considerations are tolerance stickups, relative motion between drive motor and pump, possibleangular displacements on installations, spline teeth dimensions, etc. Improper alignment can cause

    excessive vibration (leading to premature failure), or failure of the pump shaft seal.

    Splines Beyond alignment, spline wear is an important consideration. Usually some lubrication(grease) is applied to the splines to minimize wear. Selection of lubrication should includetemperature, corrosion inhibiting and reasonable life of the grease before breakdown occurs.

    Torque Requirements Both start-up torque and running torque should be considered. Start-uptorque is higher than running torque. Start-up torque accelerates the mass/inertia of the pump andfluid, leading to temporary high stresses within the mounting hardware and pump. This is more of aconcern on APU and RAT installations. Obviously, the speed of the drive motor must match themanufacturers recommended speed for the pump (usually accomplished through a gear box).

    Axial and Radial Shaft Load Capability Ensures pump shaft and splines are adequately sized for

    static and fatigue loads that the pump will see over its operating life.

    Case Drain Line A case drain line is installed to drain pump leakage flow back to the reservoir. Casedrain back pressures affect seals and bearings (via load balance across them), balance and loading ofthe pump rotating hardware, and piston leakage characteristics. In most aircraft installations, the casedrain line back pressure is equal to the reservoir pressure (which is approximately the pump inletpressure). This minimizes leakage on the intake side of the pump. Normally, back pressure from thecase drain line is not an issue, but if back pressures are abnormally high (such as clogged filter), theeffects on the pump should be looked at more closely. Case drain pressure < 150 psi is a rule ofthumb for good pump life. The case drain line must be large enough to cover the maximum case drainflow at the nominal case drain pressure or even a slightly lower pressure. The case drain line willnormally flow back to return through the return line filters. In some applications, a separate filter isused on the case drain line.

    Inlet LineThe inlet line to the pump is designed as part of the pump inlet system. The pump inletsystem consist of reservoir and reservoir pressurization and tubes/hoses from the pressure to thepump. This system should be designed to ensure fluid is provided to the pump inlet within the inletpressure range and viscosity recommended by the manufacturer. In sizing the inlet line, the length ofthe tubing, bends, height fluid is pumped, reservoir pressurization, additional components (such as aheat exchanger) in the line and other factors should be taken into account. Sizing of a pump inlet lineuses basic pipe flow equations and reservoir (or supply) pressurization (see Reservoir, HydraulicDescription).

    Outlet Line The pump outlet line should be sized to system pressure drop requirements and tominimize affects of pump pressure pulsations. Primary considerations in design of the outlet line arepump pressure pulsations, accumulators (see Accumulator, Hydraulic - Description), pump system

    response and parallel pump installation.

    Regarding pressure fluctuations, hydraulic fluid has mass and is compressible. Hence the oil in thepump downstream tubing behaves like a very stiff spring, with variable stiffness as the downstreamconfiguration changes. Pulsations are a result of each piston within the pump transferring a discreteamount of fluid to the system, leading to a pulsed input in the hydraulic system. The flow pulses decayover time from the damping provided by internal flow friction in the downstream tubing andcomponents. The pulsation frequencies for a odd numbered piston pump are

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    Example: For a pump running at 2700 rpm with 9 pistons

    For aircraft systems, both pulsation frequencies are usually above the response frequency of thedownstream components, however, in some cases the effects may need to be analyzed. The lowerfrequency is usually more dominant with pump noise, but both should be analyzed.

    System Interaction

    System interaction occurs when the natural frequency of the pump compensator is at or near thenatural frequency of a downstream component (such as a servo valve or actuator). Generally there issufficient difference between the natural frequencies so that system interactions do not occur.

    Another source of interaction occurs when pumps are connected in a parallel arrangement. This

    interaction can be stopped by installation of check valves in the outlet lines of each pump.

    Pump Pulsation Damping

    Several methods exist to dampen the effect of pressure pulsations from a pump:

    1. Change configuration (geometry, parts, characteristics, etc.)2. Increase volumes in pump outlet line3. Install accumulator close to the pump. Some research shows that for the accumulator to be effective, it

    should be installed with 0.3 meters of the pump, and the supply line between the main line and

    accumulator should be between 5 and 10 centimeters in length. Also, the volume of the gas accumulatorshould be sufficient so that its resonance (response) frequency is less than the pump pulsationfrequency.

    4. Install a hose at the pump outlet, or downstream plumbing.5. Install a Helmholtz resonator (H-filter) in the pump outlet line. A H-filter consists of two lines in

    series, of different volumes, that branch away from the main line. By properly selecting thelengths and cross-sectioal area of both lines, the H-filter can be tuned to a specific frequency.

    6. Install a Quincke tube in the main line. The Quincke tube is a side line with areas based on mainline area and lengths sized for a specific frequency.

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    Accumulator, Hydraulic - Description

    Accumulators store hydraulic energy and then provide this energy back to the system when required.Accumulators store energy when hydraulic system pressure is greater the accumulator and providehydraulic energy when the accumulator pressure is greater than the system pressure. By storing andproviding hydraulic energy, accumulators can perform 5 basic functions for hydraulic systems:

    Supply oil for high transient flow demands when pump cant keep up

    Help reduce pump ripple and pressure transients

    Absorb hydraulic shock waves (due to valve closures or actuators hitting stops)

    Used as a primary power source for small (low demand) systems

    Help system accommodate thermal expansion of the fluid

    Almost all aerospace hydraulic systems use accumulators for one of the above reasons. In fact, most

    hydraulic systems use an accumulator to dampen pressure transients in the power generation system.The pressure transients result from pump ripple, opening/closing of valves, actuators bottoming outand so on. Some practitioners believe accumulators are over utilized and systems can be designedwithout an accumulator in the power generation system. However, this has not been standard practiceand if an accumulator is not used, other design considerations should be considered. The selectionand design characteristics of accumulators will vary between the applications.

    Hydraulic accumulators store hydraulic fluid under pressure. Pressure is supplied through a bag,diaphragm or piston by either a spring, or pressured gas (most common). Accumulators are inherentlydynamic devices they function when configuration changes (actuators moving, valves opening, etc.)are occurring within a hydraulic system. Accumulators respond very fast to configuration changes,nearly instanteously for gas accumulators. The capability and affect of the accumulator is determinedby the overall volume of the accumulator and preload/precharge of the spring/gas

    Gas accumulators take advantage of the fact that the gas (nitrogen) is compressible. A gasaccumulator has a gas precharge that is less than nominal hydraulic system pressure. As hydraulicfluid enters the accumulator the gas is compressed to the nominal system pressure, which is anequilibrium position and represents the maximum amount of energy stored by the accumulator. Assystem hydraulic pressure drops, the gas will expand pushing hydraulic fluid back into the system.The gas precharge level is an important parameter for gas accumulators since the precharge andoverall accumulator volume determine the maximum amount of hydraulic energy that will be availableto the system.

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    There are 4 types of accumulators: bladder, diaphragm bladder, piston (either spring or gascontrolled) and metal bellows. The choice of accumulator to use in a given application depends onrequired speed of accumulator response, weight, reliability and cost. Pressurized gas accumulatorswill have the faster dynamic response and are reliable. Metal bellows accumulators are very reliable,but will not respond as fast as a pressurized gas accumulator. Accumulators with seals generally havethe lowest reliability.

    Accumulators are either spherical or cylindrical in design. Bag, piston and metal bellows accumulatorsare cylindrical. Diaphragm accumulators may be spherical or cylindrical. Accumulators are usuallymanufactured into 2 halves which are either welded or threaded together. A fill port is installed at oneend of a gas accumulator and the hydraulic connection fitting (with poppet valve, if required) isinstalled at the opposite end. For a spring accumulator, the non pressure side usually has a fitting thatconnects to the hydraulic reservoir (for seal leakage and to alleviate back pressure on a piston).Materials are usually steel, but accumulators may also be made from aluminum or a composite(filament wound) material.

    Bladder Accumulator

    A bladder accumulator consists of pressure vessel with an internal elastomeric bladder with

    pressurized nitrogen on one side and hydraulic fluid on the other side (system side). Figure 1 shows abladder accumulator with the 3 stages of operation, plus an overexpanded bag schematic. Theaccumulator is charged with nitrogen through a valve installed in the top. The accumulator will beprecharged to nominal pressure when the pumps are not operating, shown in Figure 1a. Whennominal hydraulic system pressure is applied the bag will be compressed to its fully compressed stateas shown in Figure 1b. When the bag is fully compressed, the nitrogen pressure and the hydraulicpressure are equal. As system pressure drops the bag expands, forcing fluid from the accumulatorinto the system as shown in Figure 1c. As the bag expands, pressure in the bag decreases. The bagwill continue to expand until the bag pressure equals the hydraulic pressure (which will be lower thannominal system pressure) or the bag fills the entire accumulator volume as shown in Figure 1d (anundesirable situation). A poppet valve keeps the bag in accumulator from being pulled into thedownstream tubing should the bag overexpand. If the bag was pulled into the downstream tubing, theaccumulator would never recharge and normal flow from the pump would be constricted. The

    maximum flow rate of the accumulator is controlled by the opening area (orifice) and the pressuredifference across the opening.

    The main advantages of a bladder accumulator are fast acting, no hysteresis, not susceptible tocontamination and consistent behavior under similar conditions. Accumulators are easy to charge withthe right equipment. Because there is no piston mass, the speed of the bladder accumulator isgoverned by the gas, which reacts very fast to changes in hydraulic system pressure. Hence bladderaccumulators are the best choice for pressure pulsation damping. Also, the bladder attachmentinternal to the accumulator has proven to be very reliable in service. Of course there is always thepotential for bladder failure, which is a failure that would not usually be detectable in service. Also,temperature differences on the gas will have some affect on performance. The main limitation ofbladder accumulators is the compression ratio (maximum system pressure to precharge pressure)which is limited to approximately 4 to 1. Hence gas accumulators will be larger than otheraccumulators for the same flow requirements. The precharge pressure is typically set to approximately80% of the minimum desired hydraulic system pressure.

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    Figure 1 Bladder Accumulator Schematic

    Diaphragm Accumulator

    A diaphragm accumulator is similar to bag accumulator except an elastomeric diaphragm is used inlieu of a bag. This would typically reduce the usable volume of the accumulator so the diaphragmaccumulator may not have volume capacity of a bladder accumulator. A schematic of a diaphragm

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    accumulator is shown in Figure 2. The behavior characteristics of a diaphragm accumulator are similarto a bag accumulator.

    Figure 2 Diaphragm Accumulator Schematic

    Diaphragm accumulators behavior will be similar to a bladder accumulator and have the sameadvantages and disadvantages. However a diaphragm accumulator may be spherical or cylindrical (orpossibly other shapes) which may be an advantage in some installations. The main difference withbladder accumulators is an increased maximum compressions ratio (maximum system pressure toprecharge pressure) of approximately 8 to 1.

    Piston Accumulator

    A gas piston accumulator is shown in Figure 3. A gas piston accumulator has a piston which slidesagainst the accumulator housing on seals. On one side of the piston is nitrogen and on the other sideis the hydraulic fluid and connection to the system. A fill port allows pressurization of the nitrogen.

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    Figure 3 Piston Accumulator Schematic

    A gas piston accumulator will not respond to transient pressures as fast as a bladder accumulator dueto the mass of the piston (frequency characteristics depend on piston mass and spring characteristicsof the nitrogen). However, a piston accumulator will have better damping due to hydraulic leakage(viscous damping) and friction between the piston and housing (coulomb friction & seal friction). Pistonaccumulators may also be more prone to leakage than other types of accumulators due to the seals.

    Piston accumulators will generally provide higher flow rates than gas accumulators for equalaccumulator volumes. This is because piston accumulators can accommodate higher pressure ratios(maximum system pressure to precharge pressure) than gas accumulators, up to 10 to 1, comparedwith bladder accumulator ratios of 4 to 1. The disadvantages of piston accumulators are that they aremore susceptible to fluid contamination, have a lower response time than bladder (unless the pistonaccumulator is at a very high pressure) and will have hysteresis from the seal friction. The prechargefor a gas piston accumulator is typically set to around 90% of minimum desired hydraulic systempressure.

    A schematic of a spring piston accumulator is shown in Figure 4.

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    Figure 4 Spring Controlled Accumulator Schematic

    In a spring accumulator, the spring applies a force to a piston which compresses (or pressurizes) thefluid in the accumulator. As normal system pressure, the spring will be fully compressed. As systemflow demands exceed the pump capacity, the spring will extend pushing the piston which in turnpushes fluid into the adjoining pipe. Hence the accumulator supplements pump flow.

    The maximum response time of the accumulator is set by the natural frequency, which is computedusing


    Metal Bellows

    Figure 5 shows a metal bellows accumulator. Metal bellows accumulators are used where a fastresponse time is not critical yet reliability is important. Emergency brake accumulators are a goodapplication for metal bellows accumulators. A metal bellows accumulator is shown in Figure 5. Themetal bellows accumulator consists of a pressure vessel with a metal bellows assembly separatingfluid and nitrogen. The accumulator is similar to a piston accumulator, except a metal bellows replacespiston and piston seals. Metal bellows accumulators are very reliable and long life components, andhave a proven service history. Metal bellows accumulators are pre-charged by supplier and thenpermanently sealed leading to a maintenance free accumulator. Metal bellows accumulators will beslow in responding to pressure changes due to increased mass of piston and bellows.

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    Figure 5 Metal Bellows Accumulator Schematic

    Gas Accummulator Precharging

    The precharge is the pressure of the gas in the accumulator without hydraulic fluid in the fluid side. Agas accumulator is precharged with nitrogen gas when there is no hydraulic fluid in the accumulator tothe desired pressure. A rule of thumb for bladder accumulators is to set the precharge pressure toapproximately 80% of the desired minimum hydraulic system pressure. A rule of thumb for gas piston

    accumulators is to set the precharge pressure to approximately 90% of the of the desired minimumhydraulic system pressure.

    The gas accumulator pre-charge is a very important variable for ensuring optimal accumulatorperformance and maintaining long life of the accumulator. Too high of a precharge pressure and thefluid volume capacity is reduced. Furthermore, if a bag accumulator charge is too high than the bagmay hit the poppet valve which could damage the bag through repeated hits in service, or cause afatigue failure in the poppet valve assembly. For a piston accumulator, the piston may be driven intothe stops repeatedly affecting seals or cause a fatigue failure in the piston stop. Too low of aprecharge pressure and the accumulator may not maintain desired minimum hydraulic systempressure. Also a low precharge pressure will allow a piston accumulator to repeatedly hit the upstops leading to premature failure of the accumulator. For a bag accumulator, the bag may be forcedinto an unnatural shape (e.g., with folds) leading to bag damage and premature bag failure.

    When sizing an accumulator the precharge pressure is an input to the sizing process. However, oncethe accumulator is sized the minimum and maximum gas volumes should be computed (under worstcase conditions) and analyzed to ensure piston stops are not hit or that a bag cannot fully collapse orexpand completely in the accumulator.

    Accumulator Design Considerations

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    The most important characteristics for hydraulic accumulators are listed below.

    Accumulator Type as described above there are 4 basic types of hydraulic accumulators (bladder,diaphragm, piston, and metal bellows). Each type has advantages and disadvantages and thespecifications will vary between the 4 types.

    Accumulator Volume total volume of the accumulator (both gas and fluid volume)

    Nominal Hydraulic System Pressure this is the nominal hydraulic system pressure in the system,which will usually be the no flow rating of the hydraulic pump

    Minimum Hydraulic System Pressure this is the minimum pressure that the accumulator mustmaintain in the hydraulic system. This is a design requirement used to size the accumulator.

    Precharge Pressure precharge is the pressure of the nitrogen in an accumulator without anyhydraulic fluid in the accumulator. The precharge pressure determines the amount of fluid that anaccumulator can hold at the system pressure and the desired minimum hydraulic system pressure.

    Required Flow Rate to maintain minimum hydraulic system pressure, the accumulator must be ableto supply sufficient flow over a determined period of time. The required flow rate is a key requirementthat drives the size of the accumulator. The accumulator volume for hydraulic flow is equal to Q * t(required flow rate times the time required for this flow). The accumulator must provide this flow whenthe gas (or spring) is between the nominal hydraulic system pressure and the minimum desiredhydraulic system pressure. Note that the flow rate provided by the accumulator will be nonlinearbecause as the gas expands the pressure drops off nonlinearly. This is a design requirement used insizing accumulators.

    Output Volume Capability the output volume capacity of the fluid volume the accumulator is capableof providing between the nominal hydraulic system pressure and the required minimum hydraulicpressure. This volume must be provided at the required flow rate (see Required Flow Rate). The valueis also called the working volume.

    Response Time this is the time for the accumulator to provide the desired fluid volume. Theresponse time times the output volume capability equals the flow rate of the accumulator. This will bea function of precharge value and the flow opening (orifice) in the accumulator.

    Recharge Time this is the time fully charge an accumulator from a fully drained state (i.e., atminimum volume, which is the volume at the minimum desired hydraulic system pressure). Thisshould be evaluated when there is a fast duty cycle requirement. The recharge time will be the amountof time for fluid to fill the accumulator based on the available flow rate from the pump (minus othersystem demands).

    High Frequency Cycling Capability only a concern when accumulators are used for damping ofpressure pulsations or very fast pressure transients. In this type of application, the frequency response

    capability of the accumulator should be computed to ensure it is compatible with the transientphenomena.

    Fluid Type accumulator seals and elastromeric bladder/diaphragm material must be compatible withthe hydraulic fluid used in the system

    Failure Modes the main failure modes for an accumulator will be failure of a bladder or piston seal,or a pressure vessel burst. The affects of a loss of accumulator performance should be evaluated inthe hydraulic system to ensure no unacceptable affects may occur within the system. For a potential

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    pressure vessel burst, the installation should be reviewed with respect to surrounding componentsand also for drainage of fluid and compartment ventilation.

    Operating Temperature Range the behavior of the gas (nitrogen) varies with temperature.Accumulator performance should be evaluated over the expected temperature range of the nitrogen.

    Mounting Position vertical is always preferred with fluid outlet at the bottom. Horizontal installationswill tend to wear a bladder or diaphragm on the down side leading to earlier failures and lowerreliability. For piston accumulators, the seals will also tend to wear unevenly leading to earlier leakage.If a non-vertical installation is required some evaluation of accumulator life should be accomplishedand the appropriate maintenance inspections (or life limits) put in place.

    Mounting Flange Determine method of mounting accumulator is acceptable in your application andthat the mount is capable of withstanding all mounting forces, including crash g loads. Analysis shoulduse the mass of the accumulator when fully charged with fluid.

    Connection Fitting the hydraulic interface fitting must be known so that a mating fitting can beincluded in the design of the hydraulic system.

    Applications of Accumulators

    One of the main applications of hydraulic accumulators is to supply flow for brief periods of time whena pump cannot keep up. A benefit of using an accumulator in this regard is that it allows the pump sizeto be smaller. Usually the accumulator only assists during a worst case duty cycle or after a particularfailure has occurred in the system. This requires having an accumulator of sufficient volume to supplythe flow needs while still maintaining adequate system pressure. The approach to sizing anaccumulator for this application is shown in the sizing section (see Accumulators, Hydraulic Sizing).

    Another application of accumulators is to damp pressure spikes from pumps or downstream

    configuration changes (such as actuators hitting stops and valves closing). This is most often done inthe power generation portion of a hydraulic system, but accumulators can be put anywhere in thesystem for pressure pulsation damping. In this application, as a pressure wave moves up and downthe piping, the energy is partially absorbed by the accumulator each time the wave flows by theaccumulator. Hence the wave damps out much faster than in a system without an accumulator.Standard practice has shown this to be a proven technique, but no well proven design procedureexists for both sizing and placement of accumulators for pressure pulsation damping. Hence someexperimentation may be required if an initial design does not achieve the desired results. One of thereasons for experimentation is that laboratory research has shown pressure waves in pipe to be botha function of time and location along a pipe. Thus at some locations along a tube there will only besmall changes in pressure magnitude (high peak to low peak) while at other locations the pressurefluctuations (high peak to low peak) will be much larger.

    A secondary function of accumulators is to absorb volume changes in fluid due to temperaturefluctuations. If an accumulator is not used and a rise in temperature increases pressure above systempressure, then the fluid must flow through a thermal relief valve to the reservoir. This is wasted flowand hence results in wasted energy thereby decreasing system efficiency. Furthermore, a pressurerelief valve exhibits hysteresis and must flow a sufficient fluid so that pressure drops below the levelwhere the valve will close (which could be a significant flow amount). With an accumulator, reasonablevolume changes can be accommodated without having flow to the reservoir. Computing accumulatorsize to accommodate temperature variations is relatively straightforward.

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    Of course, an accumulator can be sized and installed to do multiple functions. Therefore, a singleaccumulator can perform any or all of the above functions. The size and type of accumulator chosenwill depend on the functions that accumulator is addressing.

    Lastly, the loss of pressurized gas in a sealed accumulator (or spring failure in spring accumulators) isgenerally a latent failure. This latency may be an issue when conducting a safety analysis on a system

    where the accumulator plays an important role (such as emergency gear extend or emergencybraking). In this case, an acceptable functional test procedure will usually need to be implemented atan appropriate interval.

    Check Valve, Hydraulic, Description

    The function of a check valve is to prevent flow in one direction andallow flow in the other direction. Check valves commonly use a poppetand light spring to control flow as shown in the figure below. If P1A1 >P2A2 + spring force + friction, then flow occurs in the direction of thearrows. If P1A1 < P2A2 + spring force + friction, then the poppet would bepushed to the left, against the stop, prohibiting flow in the reversedirection.

    Figure 1 Simplified Check Valve Schematic

    The most common method for designing a check valve is illustrated inthe Figure 1. Different manufacturers may utilize other designapproaches. For example, another type of check valve is a ball thatpushes against a spring. Operation is similar to the check valve shownin Figure 1 except a ball replaces the piston.

    Check valves are used in hydraulic systems anytime flow in a selecteddirection is not desirable or may create a problem. Check valves are notused in bidirectional flow lines, such as to and from actuators. Someexamples where check valves are used are

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    Return lines - to prevent a leakupstream of the reservoir from drainingthe reservoir

    Return lines to prevent a pressurespike from migrating back up a return

    line to a component (this is especiallyimportant where actuators havemechanical locks and a pressure spikein the return line could cause the lock todisengage)

    Ground Service lines to ensure flowonly flows in the proper direction whenservicing and also to prevent a groundservice line leak from draining thesystem

    Pressure lines to ensure no reverse

    flow through a hydraulic motor

    Low Pressure (Feed) Line to a Pumpto prevent fluid in the pump fromflowing back to the reservoir withoutgoing through a filter

    Charged Accumulator Lines - checkvalves can be used to trap pressure ina given volume and maintain thecharge pressure for a specified timeinterval. An example of this applicationwould be using a pump to charge an

    emergency brake accumulator througha check valve. In this case, when thepump is turned off pressure ismaintained downstream of the checkvalve and is available for emergencybraking (see Figure 2).

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    Figure 2 Charged Brake Accumulator using Check Valve

    Power Control Units (PCUs) toprevent backflow out of a PCUs supplypressure port should the supplypressure to that PCU fail. This isusually required to ensure hydraulicfluid is maintained in the PCU in orderto prevent lost of the actuator stiffness.

    This is generally required on PCUs thatdrive flight control surfaces where flutteris of concern. The inlet check valve isgenerally used with a return linepressure compensator that will maintaina low pressure (50-150 psi) in the PCUinternal porting for a given amount oftime following the removal of the inletsupply pressure to the PCU.

    Parallel Pump Installations to preventoutlet flow from one pump at a slightlyhigher pressure from flowing to the

    other pump (see Figure 3). In this typeof pump installation, one of the twocheck valves is designed to open at aslightly higher pressure than the other.For example, say the left pump ispowered by the aircrafts engine andthe right pump is power by an electricmotor. In this type of application theengine driven pump generally has alarger flow capacity over that of the

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    electrical driven pump. The electricaldriven pump is intended as a backupfor the engine driven pump. In this casethe check valve on the electrical drivenpump could be set to open (crack)about 50 psi more than the enginedriven pump. This will ensure theengine driven pump supplies thesystem leakage flow during periods ofno hydraulic flow demands and will alsohelp reduce electrical motor noisevariability.

    Figure 3 Use of Check Valves in a Parallel Pump Arrangement

    When considering the use of a check valve, the following factors should

    be evaluated:

    Pressure Rating make sure the valve is rated appropriately for thesystem pressure

    Regulation Range what is the minimum and maximum prequired to go into and out of the checked flow position

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    Pressure Drop Across the Valve this will affect design pressureavailable to downstream components and thus, the sizing of thosecomponents

    Temperature Rating valve should be rated for fluid temperatures andapplicable environmental temperatures

    Valve Materials valve should be sufficient to pass proof and bursttesting, not be susceptible to corrosion and other environmentalconsiderations, operate properly under temperature extremes

    Seals/Clearances affects overall reliability of the valve. Some valvesmay not use seals and will maintain tight clearances between piston andhousing to minimize leakage through the valve. The designcharacteristics can be affected by environmental conditions andaging/wear over time.

    Leakage what is the leakage through the valve in the checked positionunder all environmental conditions?

    Failure Modes the dominant failure modes consist of the pistonjamming in either the open or the closed position. Clogging is also apossibility.

    Chattering valve should be evaluated for potential to exhibit chatteringor limit cycle behavior under certain upstream or downstreamconditions. This will be a function of the natural frequency of the servoand the damping.

    Directional Control Valve, Hydraulic - Description

    Hydraulic valves function to control pressure, control flow or direct flow in response to externalcommands. Directional valves are valves that direct flow in response to external commands.Directional valves are usually servos (see servos) where the servo is positioned in response tosolenoids, torque motors or mechanical input. Directional valves do not provide flow or pressureregulation and functional only to direct flow (much like a switch). Sometimes directional valves arepackaged with other components such as orifices or check valves. This has the advantage ofcombining several functions into 1 assembly (or 1 part number) to simplify installation. Directionalvalves are either open or closed (in 1 position or another). Directional valves do not utilize a spool in asleeve design, but quite frequently use this configuration. Also, the spool may be zero lapped or haveunderlapping or overlapping designs.

    Directional valves are referenced by the number of positions the spool will take (2, 3 or 4 positions aretypical) and the number of hydraulic ports in the valve (2 way, 3 way and 4 way are typical). Examplesare shown below.

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    Two Way, Two Position (2/2) Valve

    In a two way, two position valve, the servo can be in one of two positions and the two ways because

    there are 2 fluid ports in the valve (or, if you prefer, the valve housing). Although a spool arrangementis shown, any type of check valve could be considered a two way, two position valve.

    (Two Positions Shown)

    Three Way, Two Position (3/2) Valve

    In a three way, two position valve, there are three inlet/outlet ports in the valve and the spool can be inone of two positions. A 3/2 valve would be used to allow fluid flow into or out of actuator or motor.

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    Four Way, Two Position (4/2) Valve

    In a four way, two position valves there are four inlet/outlet ports in the valve and the spool can be located in one oftwo positions. For 4/2 valve fluid is always flowing through the valve with system pressure supplied to one of the twooutlet ports at all times. The other port would then be ported to return. 4/2 valves would normally be used in hydraulicsystems in conjunction with an upstream shut valve (or 2/2 valve). In this case a 4/3 valve usually makes more sense.However, 4/3 valves can be found in power control units (PCUs), where a shutoff valve is installed in the PCU wherea shut valve is not packaged with the 4/2 valve due to other design considerations in the PCU.

    Four Way, Three Position (4/3) Valve

    In a four way, three position valve, the spool is in one of three positions and there are 4 inlet/outletports in the valve. In the midstroke position there is no flow through the valve. A good application of a4/3 valve is actuator control, where the actuator control goal is to extend, retract or hold a position. 4/3valves are used in servovalves, where the spool is controlled by a flapper valve or a jet pipe valve.

    When specifying a directional control valve, the following parameters should be evaluated:

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    Pressure Rating make sure valve is rated for your system pressure

    Pressure Dropthis is the manufacturers pressure drop at a rated flow through the valve. There maybe a tolerance on the pressure drop which may need to be evaluated

    Pressure Drop Variance how does the pressure drop change for non-rated flow conditions?

    Flow Control A directional flow may also incorporate flow control. This would normally beaccomplished through port sizing or putting an orifice in the outlet flow. If flow control is part of thevalve, then the method of flow control should be ascertained and the accuracy and tolerance of theflow for design and off-design conditions should be evaluated.

    Temperature Rating valve should be rated for fluid temperatures and applicable environmentaltemperatures

    Actuation Time time to move from open to closed may be important in some applications

    Valve Materials should be sufficient to pass proof and burst testing, not be susceptible to corrosionand other environmental considerations, and not cause any problems under temperature extremes

    Seals/Clearances affects overall reliability of the valve. Some valves may not use seals and willmaintain tight clearances between spool and housing to minimize leakage across the servo pistons.The design characteristics can be affected by environmental conditions and aging/wear over time.See Seals - HydraulicComponents for discussion on seals.

    Method of Spool Position Control the directional valve can be controlled by a solenoid, a linkage, atorque motor, a pneumatic source or a hydraulic pressure source. This source needs to be evaluatedunder all foreseeable conditions to ensure if will open and close the valve as required. Additionalspecifications will likely be required for the control element.

    Leakage does the valve have high or significant leakage levels? Leakage is wasted energy.

    Chattering a directional valve should be evaluated for potential to exhibit chattering or limit cyclebehavior under certain upstream or downstream conditions. This will be a function of the naturalfrequency of the servo as well as the damping and friction levels. See Friction - HydraulicComponents for further discussion of friction characteristics.

    Failure Modes the main failure mode jam in any position from full closed to full open andcontamination.

    Flow Characteristics

    Flow ratings (pressure drop for a rated flow) for directional valves are normally provided by themanufacturer. Directional valves may incorporate provide flow control through either outlet port sizingor an installation of an orifice in the valve outlet. Flow behavior would follow the orifice flow equation(see Orifice Flow Hydraulic). For the affects of valve lapping on valve flow see servo section.

    Component Qualification

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    See Qualification - Hydraulic Components for discussion on directional valve qualification and requiredcertification testing.

    Pressure and Flow Characteristics for a 4/3 Valve

    Reference the 4/3 directional schematic above, note the flow areas from Ps to PA and PB to Pr areequal (matched and symmetrical valve). Ignoring leakage through the servo piston, the flow rates arecharacterized by the orifice flow equations,



    The flow areas A1 and A2 are functions of valve position, xv. The load pressure drop and load flow aregiven by


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    Since Q1 = Q2, equations (1) and (2) can be combined, such that


    Since return pressure is small compared to system pressure, we can let P r = 0, leading to equation (6)


    Equations (3) and (6) can be combined to obtain



    Equations (7) and (8) relate PA and PB to supply pressure and load pressure. If the load pressure iszero, PA and PB are equal to of the supply pressure.

    For the load flow rate, equations (7) and (8) along with equations (1) and (2) can be substituted intoequation (4) to yield an equation for load flow in terms of supply pressure and load pressure. The finalequation is


    This equation assumes Pr is negligible. Also, note that A1 = A2.

    The above relationships for PL and QL are important for understanding theoretical pressure flowcharacteristics of a four way, three position valve where the servo can be positioned in any positionalong the bushing (infinitely variable flow area). These equations are used is the discussion ofhydraulic servovalves (seeServovalve, Hydraulic Description).

    Filters, Hydraulic - Description

    Filters are necessary in hydraulic systems for filtering out contamination and debris. Contaminationand debris in hydraulic systems come from many sources. Contamination includes metal flakes, glass,

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    ash, lint, various fibers, rubber, sand, etc. Contamination is very small generally in the 1 100micron range (1 micron = 1.0E-06 meter). Over time contamination can impact component reliabilityand performance of components. The purpose of hydraulic filters is to remove contamination from asystem and keep the fluid cleanliness within design tolerances.

    During the formation of the hydraulic oil, the fluid will not be pure and contamination will be present.

    Therefore, when put into the system there will be some amount of contamination in the fluid. This typeof contamination includes dust and ash plus small particulates/residues from processing equipment.Small amounts of water may also be present in the fluid. During storage and shipping, dust, paint,chips of various materials and possibly paint can be introduced into the fluid.

    Once installed in a vehicle, the main sources of contamination is contamination that exists incomponents and tubes (from manufacture or introduced during installation) and wear that occursduring normal operation. During manufacture and assembly a common practice is to flushcomponents to remove any debris arising during the manufacturing process. On airplanes, fluid l inescan be flushed prior to connection of the hydraulic components. This is also a common practice. Thegreatest source of in-service contamination is the hydraulic pump. Pumps operate at high speeds in asevere environment and contamination (metal filings and seal pieces) from the rotating componentsoccurs in every pump. This contamination makes its way into the system through the case drain line,

    which is why case drain fluid must go through its own filter or through the return filter. Should a pumplose inlet fluid the pump runs dry and contamination from the pump increases dramatically. Normallyafter a failure of this nature, all filters are replaced in a system. Other components with moving partswill also create contamination, but usually at a rate less than a pump. Another source of contaminationis hydraulic ground power carts. Sometimes carts are not well maintained or the filter on the cart is notto the level required by the airplane. This is why ground service connections always flow directly to thefilters before entering the system.

    The overall fluid cleanliness for a hydraulic system is defined by a Class number. Class numbers runfrom 1 to 12. The Class number is determined by the number and size of particulates (contamination)in the fluid. Class number is determined through a laboratory analysis of a representative fluid samplefrom the system. The lower the Class number the cleaner the fluid. Aerospace vehicles are usually inthe 7 to 9 class range. Table 1 shows the maximum amount of particles for a given particle diameter

    for Class 6, 7, & 8 systems. As shown in the table, Class 7 system will have 32,000 particles/100 mLin the 5 to 15 m range, 5,700 particles/100 mL in the 15 to 25 m range, 1012 particles/100 mL inthe 25 to 50 m range, 180 particles/100 mL in the 50 to 100 m range, and 32 particles/100 mL in> 100 m range. Going to Class 6 divides the numberof particulates by a factor of 2. Going to Class 5halves the allowable particulate count from the Class 6 level (or divides by 4 from the Class 7 levels).Similarly a Class 8 system is allowed twice as many particulates in each size range as a Class 7.



    Class 6 Class 7 Class 8

    5 to 15m

    16,000 particles/100mL

    32,000 particles/100mL

    64,000 particles/100mL

    15 to 25m

    2,850 particles/100mL

    5,700 particles/100mL

    11,400 particles/100mL

    25 to 50m

    506 particles/100 mL 1012 particles/100mL

    2024 particles/100mL

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    50 to 100m

    90 particles/100 mL 180 particles/100 mL 360 particles/100 mL

    > 100 m 16 particles/100 mL 32 particles/100 mL 64 particles/100 mL

    Table 1 Particle Size for Class Rating

    A simple filter would consist of fine mesh screen or more appropriately, a number of fine meshscreens put in series so that the fluid has to flow though many meshes. Other methods to filtercontaminates would be a membrane, woven wire cloth element, synthetic fiber, cellulose, micro-glassand metal edge element. A mesh filter is shown in Figure 1. In this example, the filter assembles into acontainer which threads into a filter manifold (not shown). A filter manifold is a machined housingcontaining pressure, return and case drain (if used) filters. Inlet and outlet ports are also included inthe housing for connecting the appropriate pressure and return tubes. The advantage of the filtermanifold is all filters are contained in one location for ease of inspection and maintenance.

    Figure 1 Mesh Filter

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    A filter schematic is shown in Figure 2. This schematic shows a common filter arrangement and theflow paths through the filter. An additional feature is shown in Figure 2. This feature is a high ppopup switch or button. During normal operation low p across the filter the delta pressure across theswitch is not sufficient to overcome the spring force and the button remains recessed. If the pressuredrop across the filter increases to a predetermined levelhigh p across the filter the deltapressure across the switch piston will overcome the spring and push the button up. This is illustratedin the enlarged portion in Figure 2. In lieu of a pop button, an electrical switch can be used. With anelectrical switch the high p piston will actuate a switch which will provide indication to a groundservice panel and/or flight compartment.

    Figure 2 Hydraulic Filter Schematic with p Pop Up Indicator

    The main considerations of a filter are filtration level, efficiency rating and flow versus pressure dropcharacteristic. Filtration capability is listed as a micron level. So a 5 micron filter will filter out particleswith a diameter or width 5 microns or greater. Particles smaller than 5 microns will be able to flowthrough the filter. Filters can be rated as 100% efficient which means that all particles equal to orgreater than the filter rating will be caught in the filter. If the rating is less than 100% than some

    particles in the filters range will make it through the filter. Efficiency ratings are sometimes given interms of a beta rating. Beta ratings are defined in the following way: For a 5 micron filter, a beta ratingof 200 means that for every 200 particles > 5 microns only 1 particle greater than 5 microns will makeit through the filter ( = 200/1 = 200) and a beta rating of 2 means that that for every 200 particles > 5microns 100 particles greater than 5 microns will make it through the filter (=200/100=2). Thus,higher beta ratings imply a more efficient filter. Filters that are 100% efficient (i.e. = ) are referred toas absolute filters.

    Since filters must inherently restrict flow there will be a pressure drop (flow resistance) though thefilter. As a general rule, as filter efficiency goes up (i.e., more and/or finer meshes) pressure drop will

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    increase. Also, as filters get dirty flow resistance goes up. When evaluating a filter in the system thepressure drop versus flow rate characteristic should be evaluated. The characteristics should beevaluated at the most dirty condition (i.e., just prior to filter indication). Hydraulic filters are oftenmanufactured as a filter manifold assembly. This means that more than one filter may be part of thefilter manifold. Also, the manifold allows easy access to the filter to support periodic filterreplacements. Old filters may be discarded. In some filters, it may be possible to send filters back tothe manufacture for cleaning and recycling.

    Filters normally have a indicator as part of the assembly (see Figure 2). The indicator will use a deltapressure device that senses p across the filter. When the p is sufficiently high the indicator pistonwill move to a position where a switch is closed or visual indication shows. The indicator alertsmaintenance personnel that the filter needs replacement. Filter indications are often checked on a pre-flight walk around by flight crew and maintenance personnel. If a switch is used, the switch can beconnected to crew compartment indication. In some bypass filter designs, an indication is providedwhen the bypass valve goes into bypass mode.

    A filter may include a filter bypass provision, as shown in Figure 3. A filter bypass allows fluid to flowaround (bypass) the filter when the p across the filter is high (i.e., filter is clogged). A bypassfunctions similar to a priority valve where the p overcomes a spring when the p is sufficiently high.

    A good practice is put bypass provisions on filters in the return line filter. Generally it is preferable tohave dirty fluid flow to the pump rather than no flow. No flow to the pump will require the pump to bereplaced and will put much more contamination into the system as the pump starts to run dry. Bypassprovisions on pressure lines is generally not required nor preferred. If a bypass provision is provided inthe pressure line filter then when the filter is clogged and goes into bypass contamination will make itsway to all downstream components. This can affect performance of many components in a systemand lead to (expensive) replacement of numerous components in a system. If a filter is used on a casedrain line, a good practice would be to install a filter with a bypass provision. This will preventexcessive case drain back pressure and premature pump failure.

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    Figure 3 Hydraulic Filter with Bypass Valve

    Another type of filter is shown in Figure 4. This filter is an inline filter and would be representative of afilter used to filter flow just upstream of a hydraulic component, such as a PCU.

    Figure 4 Inline Filter

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    Filter requirements come from component cleanliness requirements. can be put on main systemfilter or a separate filter can be put in the lines flowing to the component or internal to the component(as is often the case with PCUs). As discussed in System Design, Hydraulic Power Generation, atable should be built that lists the cleanliness requirements for each component. The worst case

    requirement should be used to select an appropriate filter.

    Tighter filter requirements will lead to a cleaner system and a lower Class rating. However, there is nostandard relationship between filter capability and Class rating. To determine Class rating for a givenhydraulic system, a fluid sample will need to be taken from the system and sent to a certified lab forevaluation. Care should be taken to not introduce additional contamination into the fluid whengathering the sample. Statistical sampling methods are recommended due to variances in vehicleoperation and lab methods. There will likely be a difference in Class rating based on the length of timethe vehicle has been in service.

    When selecting a filter, the following factors should be evaluated:

    Maximum Capacity Rating capacity rating is the maximum flow that can be accommodated by thefilter at the rated pressure drop

    Maximum Pressure Rating pressure rating is the maximum pressure that will be seen by the filter inservice. Return pressure spikes should be evaluated when determining the maximum pressurerequirement. Proof and burst pressures will be based on the maximum pressure requirement.

    Flow vs.p Characteristicsfilter manufacturers can provide a flow vs. p curve for the filter. Thiscurve should be evaluated to ensure that the filter will not restrict flow up to the maximum p (cloggedfilter) indication. If a simulation model for the return system has been constructed this curve should beused for the filter model (see also Filter, Hydraulic Sizing).

    Maximum Temperature filter should be rated for the maximum combined fluid and environmentaltemperatures expected in service. This will normally be determined by the seal material used in thefilter.

    Minimum Particle Size Rating what is the filter cleanliness capability in microns (e.g., 5 micronfilter)?

    Efficiencywhat is the efficiency rating of the filter, in terms of a rating or whether is absolute (100%efficient)?

    Filter Element (Cartridge) Type will filter be wire mesh, membrane, woven wire cloth element,synthetic fiber, cellulose, micro-glass or metal edge element? What are specific maintainability issues

    associated with available filter types? Is the filter compatible with your fluid? What is the expected lifeof the filter? The pros and cons of available choices in your application should be reviewed with thefilter manufacturer.

    Reusable Element is the selected filter element reusable after cleaning or is the filter disposable? Ifreusable, what is the method for cleaning?

    Container (Housing) what is the material used for the housing? Is the material corrosion resistance?

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    Clogged Filter Indicationwill a high p (clogged filter) indicator be used? If so, what are the designcharacteristics? What type of indication is provided manual pop up button or electrical switch?

    Bypass Provision will the filter contain a bypass provision? If so, what are the design features of thebypass valve? Does the bypass valve provide indication independent of a clogged filter indicator?

    Seal Type & Material what type of seal is used in the filter? Is the seal material compatible with thefluid and operating temperatures of the fluid?

    Fittings the inlet and outlet tubing fittings need to be compatible with the fittings on the filter.Normally, these are threaded fittings.

    Mounting the method of mounting or supporting the filter within the system should be robust and notallow any loads to be imparted to the connecting tubes.

    Filter Qualification

    See Qualification - Hydraulic Components for discussion on hydraulic filter qualification and requiredcertification testing. Temperature, proof and burst pressure, vibration, impulse pressure andendurance testing would be important.

    Flow Control Valve, Hydraulic - Description

    The two methods of controlling flow rate in a hydraulic circuit are (i) using a fixed orifice and (ii) using aflow control valve.

    For accurate flow control, a device that regulates to a p across an orifice is required referred to aspressure compensated flow control valve. Figure 1 shows a simplified pressure compensated flowcontrol valve.

    Figure 1 Simplified Schematic of a Inlet Pressure Regulated Flow Control Valve

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    In this valve, the p across the flow metering orifice is maintained at a constant value producing a

    constant flow rate. The general orifice flow equation indicates that holding the orifice area and pconstant, where the fluid properties (bulk modulus and density) are relatively constant, will yield a

    constant flow rate through the orifice. To hold p constant, the upstream and downstream pressures

    are ported to different sides of a servo (piston). An adjustable spring assists the lower pressure to holdthe valve open at low input pressures. As thep (force balance) across the servo varies the flowopening by the servo, the metering orifice inlet pressure is regulated. Hence as P1 increases (orP2decreases), the servo moves to the left and reduces the servo flow area. And as P1 reduces (orP2 increases), the servo moves to the right thereby increasing the flow area. The flow control valveshown in Figure 1 modulates P1 to control flow. Calibration of the flow control valve is obtained byadjusting the metering orifice. The spring preload may also be adjustable.

    A second example of the flow control valve is shown in Figure 2. In this valve, the servo modulates P2.However, the overall function of the valve is similar to the valve in Figure 1. Regulating P2 may be anadvantage over regulating P1 if the servo port in Figure 1 could become the controlling orifice (flowarea becomes smaller than the metering orifice). In this case the servo port opening would becontrolling flow.

    Figure 2 Simplified Schematic of a Outlet Pressure Regulated Flow Control Valve

    Fixed Orifice vs Flow Control Valves

    A comparison between the use of orifices and pressure compensated flow control valves is shown in

    Figure 3. The orifice flow varies dependent on (p). The amount of variation seen in a practical

    application of orifice flow depends on the range of p seen over the operating range of the orifice in

    the system. As shown in Figure 3, the flow control valve holds flow constant over a wide range of p

    (from min regulation p to maximum p).

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    Figure 3 Notional Graph Showing Flow Characteristics for an Orifice and Flow Control Valve

    For an orifice (see Figure 4), flow is governed by the orifice flow equation (seeOrifice Flow - Hydraulic)


    As can be seen in Equation (1), for a fixed p flow can be controlled by controlling area. In practice,P1 and P2 are never constant and therefore orifices do not provide constant flow rates over alloperating conditions (system pressure, downstream pressures, temperatures, etc.). Nevertheless, inmany applications, fixed orifices can be sized to limit flow under the worst case condition and theaccuracy of a simple orifice is sufficient. As an example, simple orifices are common in landing gearactuator circuits, where the time to retract or extend the gear can be in the range of 6-10 seconds. Inthis case, an orifice can be sized to maintain the 6-10 second requirement under all operatingconditions.

    In Figure 3, note the flow rate for a flow control valve is constant over a wide pressure range.However, as the input pressure range varies so will the output pressure. This will affect the inletpressure to a downstream component. So, while the flow to a component (actuator or motor)downstream of the flow control valve will be constant, the inlet pressure to the component will change.

    This will affect the p across the component and hence the power output of the component (Power

    = p x Q).

    Sharp Edge Orifice

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    Short Tube Orifice

    Figure 4 Sharp Edge and Short Tube Orifices

    When considering the use of a flow control valve, the following factors should be evaluated

    Pressure Rating ensure valve is rated for your system pressure

    Regulated Flow Rating should be in the range desired for the specific application

    Regulation Range What is the minimum and maximum p required across the orifice for regulation?

    For example, one flow control valve manufacturer has a maximum p of 3000 psi and a minimum pof 100 psi across the metering orifice to maintain flow regulation within the flow tolerance.

    Flow Tolerance Ensure tolerance on flow regulation is sufficient for your application. For example,

    some valves will control flow to within 10% of the setting. Tolerances result primarily from frictionforces on the servo. When using flow control valves in parallel systems, such as thrust reversers orground spoiler systems, flow tolerances should be sufficiently tight to avoid undesirable asymmetricoperation between the two actuators (this may also be affected by piping differences lengths &bends - between the two actuators and source manifold).

    Pressure Drop Across the Valve in the Regulation Range This will affect design pressure availableto a downstream component and will affect sizing of that component. Keep in mind that while flow isconstant over a wide pressure range, the outlet pressure of the flow control valve can varysignificantly. This will affect available power to the downstream component.

    Temperature Rating valve should be rated for fluid temperatures and applicable environmentaltemperatures

    Valve Materials valve material(s) should be sufficient to pass proof and burst testing, not besusceptible to corrosion and other environmental considerations, and operate properly undertemperature extremes

    Seals/Clearances affects overall reliability of the valve. Some valves may not use seals and willmaintain tight clearances between spool and housing to minimize leakage across the servo pistons.The design characteristics can be affected by environmental conditions and aging/wear over time.See Seals - HydraulicComponents for discussion on seals.

    Failure Modes the dominant failure modes in the flow control consist of the servo valve jamming inany position from full closed to full open and degraded performance due to contamination. It may alsobe possible for the adjustable device on the metering orifice to fall out of adjustment leading to loweror higher flow regulation settings.

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    Chattering valve should be evaluated for potential to exhibit chattering or limit cycle behavior undercertain upstream or downstream conditions. This will be a function of the natural frequency of theservo as well as the damping and friction levels. See Friction - Hydraulic Components for furtherdiscussion of friction characteristics.

    Hysteresis how does the flow regulation change when approaching a control point from a low

    pressure condition or high pressure condition? Hysteresis affects flow control accuracy andcontributes to chattering.

    Flow Control Valve Qualification

    See Qualification - Hydraulic Components for discussion on control valve qualification and requiredcertification testing.

    Hydraulic Fluid - Properties

    Fluids, either liquids or gases, are bodies without shape. Fluids experience great changes in shape under appliedforces unless constrained in some manner. The behavior of fluids is characterized by the fluid properties listedbelow.

    Fluid Pressure Normal tension on the surface element of a fluid

    (psi) (1)

    Fluid Density Fluid mass per unit volume. Density is a function of both pressure and temperature.

    (lbm/ft3) or (lbf- sec

    2/ in


    Viscosity - During the movement of fluid, there is a tangential force that resists movement, called viscosity.

    Referring to the above figure, suppose 2 fluid layers are moving at a distance apart of dy, at a relative velocity ofdvx. Shear stress occurs between the fluid layers and is given by

    (psi) (2)

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    is the proportional constant called dynamic viscosity, with units lbf-sec/in2called reyn. is also called absolute

    viscosity or the coefficient of viscosity. varies significantly with the type of fluid and fluid temperature. alsovaries with pressure but the effects are smaller.

    Kinematic Viscosity - The ratio of coefficient of dynamic viscosity to fluid density is called the coefficient ofkinematic viscosity

    (in2/sec) (3)

    DensityThe density, , of a fluid is the mass per unit volume

    (lbm/ft3) or (lbf - sec

    2/ in

    4) (4)

    Density is a function of both pressure and temperature

    Bulk Modulus - Expanding the above equation for density in a Taylor Series (for 2 variables)


    Mass density increases as pressure increases and decreases as temperature increases as the sign of

    is positive and the sign of is negative.

    Assuming constant temperature so that = 0, then equation (5) becomes


    The quantity

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    (psi) (7)

    is the change in pressure divided by a fractional change in volume at constant temperature. The minus sign

    indicates a volume decrease with pressure increase. is called the isothermal bulk modulus (or simply bulkmodulus) since it was derived assuming constant fluid temperature. The bulk modulus represents fluidcompressibility and has a significant effect on the performance of hydraulic systems. Effects of pressure on bulkmodulus are large and effects of temperature are usually negligible.

    Effective Bulk Modulus - Both entrained air in the fluid or mechanical compliance of tubing/hoses can substantiallylower the bulk modulus. Effects of both are additive.

    Effects of Entrained Air:

    For liquid-air mixtures, an empirical formula for the effective bulk modulus, , is



    isen isentropic bulk modulus of the fluid w/o entrained air

    VG0 volume of gas entrained in the liquid at atmospheric pressure

    VL0 volume of the liquid at atmospheric pressure

    p0 atmospheric pressure

    p fluid pressure

    k isentropic exponent (normally, k=1.4)

    rv air to liquid volume ratio

    Using equation (8), the effects of entrained air are shown in Figure 1 below.

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    Figure 1 Effects of Entrained Air on Bulk Modulus

    Effects of Mechanical Compliance:

    For cylindrical pipelines, the effective bulk modulus, , can computed using


    p is the bulk modulus of the pipeline (available in Materials or Engineering Handbooks) and wis given by

    for (thick walls) (10)

    for (thin walls) (11)


    do outer pipe diameter

    di inner pipe diameter

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    Poissons number (0.3 for steel)

    S pipe wall thickness

    Fluid Properties Example: For a steel pipe with isen= 200000 psi, 0.5 O.D., S = 0.012, compute the effects ofmechanical compliance on bulk modulus.

    p = ratio of normal stress (on all faces of a cube) to a change in volume

    thin walled

    Using equation (9)

    In this example, the effects of thin walled pipe reduce bulk modulus by 25%.

    Combined Effects of Air and Mechanical Compliance:

    Effects of entrained air and mechanical compliance combine together like springs in series, i.e.,



    effective bulk modulus (psi)

    l bulk modulus of the hydraulic fluid

    MC bulk modulus for mechanical compliance

    g bulk modulus of air

    Vg volume of air in the fluid

    Vtotal total volume of fluid and air

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    Note the effective bulk modulus will be less than any of the individual bulk modulus ( l, MC or Vtg/Vg). For air,the bulk modulus is computed using

    Basic Effects of Fluid Properties

    The effects of temperature and pressure on hydraulic system fluid properties and flow characteristics are listedbelow.

    Density Effects orifice and valve volume flow rates. As density increases, orifice and valve flow rates willdecrease (see orifice flow equations).

    Increasing pressure increases density Increasing temperature decreases density

    Kinematic Viscosity Effects pipe (tube) volumetric flow rate. As viscosity increases, pipe flow rate will decrease(see orifice flow equations). Kinematic viscosity increases with increased pressure and decreasing temperature.

    Increasing pressure increases kinematic viscosity

    Increasing temperature decreases kinematic viscosity

    Bulk Modulus Effects compressibility of fluid and system response time (see pressure derivative equation). Asbulk modulus decreases, the pressure derivative will decrease leading to slower response times. Compressibilitywill affect the performance of actuators, motors and pumps because the stiffness of the fluid is less as bulkmodulus is reduced.

    Increasing pressure increases bulk modulus

    Increasing temperature decreases bulk modulus Entrained air and compliance of hoses/tubes/parts decreases bulk modulus

    Fluid Properties for Standard Hydraulic Fluids


    -54 oC -40 oC 40 oC 100 oC

    Viscosity (centistokes) 2500 600 13.2 4.9

    Bulk Modulus at 40oC and 4000 psi: 200,000 psi (minimum)

    Specific Gravitity at 60oF Approx 0.88*

    Nominal Density: 50 lbm/ft3

    Pour Point: -60oC

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    Flash Point: 82oC

    Coefficient of Thermal Expansion: 8.6E-04 cm3/ (cm


    * Not specified in MIL-PRF-5606


    -40 oC 40 oC 100 oC

    Viscosity (centistokes) 550 max 6.7 min 2.0 min

    Bulk Modulus at 40oC and 4000 psi: 200,000 psi (minimum)

    Specific Gravitity at 60oF Approx 0.88*

    Nominal Density: 49 lbm/ft3

    Pour Point: -60oC

    Flash Point: 160oC minimum

    Coefficient of Thermal Expansion: 8.2E-04 cm3/ (cm


    * Not specified in MIL-PRF-87257


    -40 oC 40 oC 105 oC 205 oC

    Viscosity (centistokes) 2200 14 3.45 1.0

    Bulk Modulus at 60oC and up to 10000 psi: 200,000 psi (minimum)

    Specific Gravitity at 60oF Approx 0.85*

    Nominal Density: 49 lbm/ft3

    Pour Point: -55oC

    Flash Point: 205o


    Coefficient of Thermal Expansion: 8.2E-04 cm3/ (cm


    AS1241 Type V (Phosphate Ester)

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    -54 oC 38 oC 99 oC

    Viscosity (centistokes) 2600


    9-12.5 3-4

    Bulk Modulus at 40oC and 4000 psi: 200,000 psi

    Density at 25oC 0.97 1.02 g/mL

    Nominal Density: 63 lbm/ft3

    Pour Point: -62oC

    Flash Point: 149oC

    Coefficient of Thermal Expansion: 1.0E-03 in3/ (in


    Motors, Hydraulic - Description

    The function of hydraulic motors is to convert hydraulic pressure and flow into rotational mechanicalenergy via an output shaft. Motors are used where a rotational output is required and actuators areused where linear output is required. However, in operation motors are more similar to pumps thanactuators and, in fact; the equations for motors are identical to the equations for pumps.

    The output of a motor is a torque and an angular velocity (note: power = torque x angular velocity).Motors operate exactly opposite of pumps and, in fact, some motors/pumps take on dual roles (willoperate as a pump or a motor depending on the position of controlling valves) in a hydraulic system.Like pumps, the governing equation for an ideal motor is


    where Q is the flow rate, p is the pressure drop across the motor, T is the output motor torque and is the motor angular velocity.

    The main parameter of interest for hydraulic motors is the displacement (similar to pumps).Displacement is the amount of fluid that is displaced (flows through) the motor for each revolution ofthe shaft, and relates motor torque to the p across motor.


    In equation (2), D is the displacement in units of in3/rev (or similar) and p = p1 p2, where p1 is the

    inlet pressure and p2 is the output pressure. The steady state performance of a pump is completelydefined by equations (1) and (2).

    There are 3 main types of hydraulic motors: piston, vane and gear.

    A piston motor can be radial or axial. In a radial piston motor, the pistons are oriented radially from theshaft. An axial piston motor is similar to an axial piston pump and is the most common motor used inaerospace. A schematic of an axial piston motor is shown in Figure 1. As can be seen in the figure,the axial piston pump is identical to a piston pump, except that the swashplate angle is now fixed (i.e.,there is no compensator and control piston). The inlet side of the motor is the high pressure side andthe outlet is low pressure. The pressure difference causes the pump to rotate. Since the swashplate is

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    fixed, speed of this motor is controlled by either controlling inlet pressure (p across the motor) or theflow rate. Also, like pumps, hydraulic motors tend to have 9 pistons, or possibly 7 (more pistonsincrease displacement and hence increase output torque). Piston motors provide the best sealing forhigh input pressures and work best in high torque, low speed applications. They have the best sealingand will be the most efficient. An axial piston motor with a fixed swashplate is unidirectional (rotate in 1direction only). To be bi-directional, the swashplate would need to be variable position. Lastly, pistonmotors will have a case drain line to allow piston leakage to flow to return.

    Figure 1 Axial Piston Hydraulic Motor

    A vane motor is shown in Figure 2. Vane motors can be a good choice for a motor in high speedapplications. The vane motor rotates as hydraulic fluid at high pressure flows through the motor to theoutlet, or low pressure side of the pump. More vanes reduce output torque ripple, but also lead tohigher pump friction. The vanes are attached to the drive shaft and fit closely to the housing ring (or

    cam ring) to minimize leakage. The vanes are pushed out by hydraulic pressure, centrifugal force orsprings (springs are shown in Figure 2). For vanes which rely on centrifugal force to extend, the vanesare may be attached to the drive shaft (or rotor) via a slot, which allows the vanes to rotate in the slotand also move slide in the radial direction. As pressure is applied, the vanes will start to rotate therotor (shaft). As speed is increased, the vanes move outward in the slots and contact the wall,providing a seal on the outer surface. The housing (or cam ring surface) can have a ramp shape tofurther reduce pressure at the pump outlet this is possible because the rotor slot allows a vane tomove radially and rotate to adjust to the housing (or cam ring) shape. For vane motors that utilizespring loaded vanes, the spring that helps to hold the vane against the housing (or cam ring) to ensuresealing at low p across the pump. This helps the motor develop starting torque faster. At higherspeeds, centrifugal force helps hold the vanes out. The vane pump shown in Figure 2 has two ports,an inlet and outlet. It is possible to have a four port vane motor, which splits the flow through twoseparate paths. A four port vane motor will have twice the torque, but will operate at the speed as a

    similar sized two port vane motor. Inlets for a four port motor will be at opposite ends of the motor tobalance bearing loads. For vane motors, a valve can control which port has high pressure and whichport has low pressure, leading to a bi-directional motor. Vane motors are not as efficient as pistonmotors but are better suited to high speed operation. Vane motors become less efficient at high inletpressures (due to potential for more slippage and leakage past the vanes). Vane motors are alsomore inefficient at low speeds. Vane motors can be operated in a reversible (bi-directional) mannerand will be less noisy then other pump types. They will generally be less expensive than pistonmotors.

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    Figure 2 Vane Hydraulic Motor

    A gear motor schematic is shown in Figure 3. This gear shows two spur gears rotating in a commonhousing. Both gears rotate, although only one gear is connected to the output shaft. Fluid enterswhere the gears mesh together as shown in Figure 3. The gears then rotate in the direction of thearrows, as the greatest pressure drop is around the outside of the housing (if the gears rotated the

    other direction then they would be pushing against system pressure). Also, by putting the input portwhere the gears mesh together puts the effective area of 2 gear teeth against the resisting pressureacting on 1 gear tooth. Gear motors work best in high speed applications. The efficiency of gear motoris lower at low speeds and increases (gets better) at high speeds.

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